WO2019038305A1

WO2019038305A1 - Control of printing presses having a plurality of main drive motors

Info

Publication number: WO2019038305A1
Application number: PCT/EP2018/072590
Authority: WO
Inventors: Nicolai Kubasiak; Eric Lenz
Original assignee: manroland sheetfed GmbH
Priority date: 2017-08-21
Filing date: 2018-08-21
Publication date: 2019-02-28
Also published as: JP7203827B2; CN111032357B; CN111032357A; DE102018120397A1; JP2020531314A; EP3672808A1

Abstract

The invention relates to the control of printing presses, in particular sheet-fed presses, having a plurality of main drive motors (M), which are connected by means of a drive train. In order to control a press rotational speed, a central control unit (ZR) is used, wherein in particular current rotational speeds of different positions of the drive train are used for the control, the rotational speeds being sensed by sensors (S) or estimated by observers. A rotational speed variable, which can be derived from a momentum equation, and/or one or more torque differences, which can be selected from the differences between the feed-in torques of the motors, are/is used as a controlled variable. The torque differences are selected in such a way that the lifting off of tooth flanks in the drive train is counteracted or prevented. A PI multi-variable controller is used as the controller structure; the controller structure can be simplified by means of an input coupling in such a way that a fixed or variable torque distribution is realized. The guidance behavior and disturbance behavior can be handled separately. The controller parameters are obtained by the optimization of a system standard, the system to be optimized having positions of important sources of disturbance as system inputs and the angle differences relevant to the printing quality as system outputs. For the positioning of the drive motors, the eigenmodes of the drive train and the positions of important sources of disturbance are used, wherein for the active torsional vibration damping the drive motors are preferably in the edge positions in the last and/or in the first module, and/or in the turning module.

Description

Control of printing machines with several main drive motors

The invention relates to a control of printing machines, in particular

Sheet-fed presses with a plurality of motors (main drive motors) according to the preamble of claim 1.

From DE 10 2008 042 396 A1 a drive for a rotary printing machine with several plants is known. The works are as printing or coating plants

educated. It is a rotary printing press with a cross-machine drive wheel train which drives a plurality of rotary bodies during a printing operation. The rotary printing press furthermore has at least one main drive motor, which feeds drive torques into the drive wheel train, wherein at least one further rotation body in the factories is assigned a single drive and the others

Rotary body with the driven by the drive wheel train rotational bodies in a torque transmitting contact. The task of this

Invention is to improve the main drive in printing presses with single drives. This is achieved in that the main drive motor is designed as a high-torque motor slow-speed, the directly or via a preferably single-stage countershaft transmission with high rigidity

Drive wheel is assigned. DE 10 2007 049 455 B4 discloses a method for operating a

Printing machine, in particular a sheet-fed offset printing press, in which at least two independently positionable drives drive in a continuous gear train of the printing machine. Coupled to the gear train to be driven facilities, in particular printing and / or coating units, the printing press are driven such that each drive to drive a respective drive associated, at least one to be driven device comprehensive subunit of the printing press is used. The Eintriebstellen the drives are rotated to provide an angular difference to each other. For the drives driving into the continuous gear train, a desired torque distribution is determined such that each drive provides a drive torque on the one hand for the subunit assigned to it and also for the subunit assigned to the respective drive downstream drive. Based on this torque distribution, an individual angle setpoint is determined for each drive. On the basis of this, each drive is individually position-controlled to an angle of rotation dependent on the torque distribution between the input points of the drives

provide. With the individual position control of the drives, the torque distribution actually occurring during the bearing control is determined.

In this case, it is checked whether the torque distribution which is actually occurring within a tolerance band defined by an upper limit value and a lower limit value lies around the desired torque distribution. If it is determined here that the actual torque distribution is within the tolerance band, the angle setpoints for the position control of the drive remain unchanged. If this determines that the actual

Torque distribution is outside the tolerance band, at least one angle setpoint for position control of a drive is automatically changed so that the actual torque distribution is back within the tolerance band.

A speed control device for a multi-motor drive for printing presses with several connected by a gear train printing units, a

Speed-controlled master motor and a torque-controlled depending on the master motor motor to another printing unit, a leading motor and the motor upstream controlled system, consisting of a

Speed sensor, which is assigned to the master motor and feeds vibration-superimposed speed values into the controlled system, a speed summation point risoii / nist and a speed controller, is known from DE 195 02 909 B4. The control system upstream of the master motor and the further motor also contains a A current regulator that distributes the total motor current in an adjustable ratio to the master motor and the motor, and to inject a speed-matched torque through the motor into the press between the current controller and the motor, an motor current limiting motor current and additional oscillatory current peaks from the motor current separating current limiter is arranged.

DE 10 2008 048 406 A1 discloses a device for controlling a

Sheet-fed rotary printing machine with at least two acting on a gear train electric drive motors known, the gear train several printing units in the sheet-fed rotary printing press mechanically

connects with each other. The invention is characterized in that the electric drive motors each have a control loop with a higher-frequency control element and that in at least one control loop of the

Drive motors a low-frequency control element is present, the

Output signal is applied in a predetermined ratio to the control loops of the drive motors.

The invention has for its object to provide a control of printing machines with multiple motors (main drive motors), with which the print quality is improved.

The object is solved by the training features of the main claim. Advantageous developments emerge from the dependent claims, the

Description and the figures.

An advantageous realization possibility is that the controlled variable for the engine speed becomes more independent of local speed fluctuations, if this is obtained by a weighted averaging of individual rpm values. Advantageously, the individual inertia of the

Powertrain can be used as weights. Another advantage is that the regulation controls the torque directions within the Drivetrain influenced and a change during operation can be prevented, whereby a possible number of edge lifting is prevented.

For example, the overall torque requirement of the printing press changes in light of changed machine conditions, such as production speed

(Engine speed), engine warming or pressure module conditions, etc., so the torque flow within the gear train changes and a changed average displacement between the individual modules adjusts itself. This pressure technology disadvantageous effect can also be counteracted with the specification of differential torques or by changing the distribution of the total drive torque to the individual main drive motors. Advantageously, the control affects not only locally on adjacent printing units, but on the entire machine and the measured speeds or the output errors have a global effect on all engines.

According to the invention, speed fluctuations are suppressed or compensated, so that a torsional vibration damping or a

Torsional vibration compensation takes place.

To control the engine speed, a central controller unit is used. In particular, several actual rotational speeds of different positions of the drive train are used for the control. The single ones

Speeds can be detected with sensors or estimated with the help of an observer. As a controller structure, a PI multi-variable regulator structure is used. This structure can be advantageously simplified with an input coupling with which a fixed or variable torque distribution can be realized. The guiding behavior (slow dynamics) and the disturbance behavior (fast dynamics) can be treated practically separately (two-step procedure).

The controlled variable used is a speed variable (machine speed) and possibly several differential torques. The

Speed variable, for example, can be advantageously derived from the pulse rate and results from the by the individual inertias weighted averaging of individual speeds or speeds. Difference torques are the differences between the

Input torques of the motors selected. The positioning of the powertrain antisubmotors will be achieved with the

Intrinsic forms of the powertrain and the positions of significant sources of interference used. In this case, in particular, the two edge positions in the last and first module, as well as in the middle region, offer themselves at the turning module. By means of suitable specification of the differential torque setpoints or a suitable choice of the distribution of the total drive torque to the individual main drive motors, a possible tooth flank lift in the drive train (gear train) can be counteracted. The controller parameters are obtained by optimizing a system standard (H2). The system to be optimized has the main fault positions as system inputs and the differential angles relevant to print quality as system outputs for calculating the system standard. By means of a regulation of printing machines, in particular

Sheet-fed presses, with one or more main drive motors, which are connected via a drive train, according to the invention for the control of a machine speed and an active damping of vibrations

(Torsional vibration damping) uses a central control unit, wherein for the control multiple speeds from different positions of the drive train are used, the speeds are detected with sensors and / or estimated with observers.

Advantageously, each speed acts on all main drive motors or multiple Ausgansfehler act on all main drive motors. Advantageously, an engine speed control variable from the

calculated weighted averages of different current speeds, the respective weights resulting from the inertia of the drive train.

Advantageously, in addition to a machine speed control variable, further controlled variables, such as preferably setting rotational torque differences, are selected. When

Controlled variable, a speed variable (engine speed) can be used, which is derived from a pulse rate, and / or one or more differential torques are used, which are selected from the differences between the feed torques of the motors.

Advantageously, a PI multivoltage controller is used as the controller structure, wherein the controller structure with an input coupling is simplified or can be coupled in such a way that a fixed or variable torque distribution dependent on a state of the printing press is realized.

Advantageously, the differential torques are selected so that a tooth flank lift is prevented in the drive train.

Advantageously, the controller design is separated according to leadership behavior and interference behavior.

Advantageously, controller parameters are optimized by optimizing a

System standard (for example, an H2 system standard) obtained, the system to be optimized, the positions of significant sources of interference as system inputs and the relevant for the print quality differential angle (for example, the

Differential angle of adjacent pressure modules) as system outputs.

Advantageously, for the positioning of the drive motors

Eigenformen the drive train and / or positions of significant sources of interference used, preferably the drive motors in the Edge positions in the last and / or in the first module, and / or are located in the area of a turning module.

Embodiments of the invention will be explained in more detail with reference to the figures, without being limited thereto. Here are shown schematically:

Fig. 1 shows a modular construction of a sheetfed offset printing press with central

controller unit

2 shows a simplified drive train model for a series oscillator, FIG. 3 normalized eigenmodes of an undamped drive carrier model

4 shows a comparison of amplitude responses with the course of the Frobenius

Norm of Η (ω),

5 shows a torque flow within the gear train,

6 shows a control structure for a PI multivalue output feedback, FIG. 7 shows a PI output feedback with an input coupling, FIG.

Fig. 8 shows the Frobenius standard for a static speed feedback, for a varied number of main drive motors with different positions on the drive train and

Fig. 9 shows the course of the Frobenius standard of a static feedback for

different system outputs with the same position of the

Main propulsion engines.

It will be the regulation of sheetfed offset presses with several

Main drive motors treated. Starting with the modeling of the

Driveline through the modal study of vibration behavior will be two different regulatory structures for several

Main propulsion engines presented. In doing so, special features specific to the press will be discussed and different drive configurations (number and position of motors) compared.

Fig. 1 shows the modular structure of a sheet-fed offset printing press. To the

The sheets are first printed on the feeder 1 by a stack isolated and successively supplied to the first printing module 3. In the first printing module 3, the single sheet is accelerated to machine speed and printed with the first printing ink. The bow goes through the following

Print modules 4 and / or paint modules 5 and is then braked on the boom 2 and stored again to a stack. Due to the modular design, a printing machine can be adapted to specific customer requirements through a variable number of modules and module combinations. For example, it is possible to provide a turning module 6 with a turning drum 7 in order to turn the sheet and to print on both sides.

The moving parts within the modules, such as rollers, cylinders, etc., are generally interconnected by a machine-gear train (not shown in FIG. 1) and form the drive train. For driving the printing machine, one or more motors M (main drive motors) may be used. An exemplary arrangement of the motors M is shown in FIG. The regulation takes place via a central control unit ZR,

Actuators M (here motors) and sensors S. The sensors S can

for example, angle or speed sensors. In this regulation, it is not absolutely necessary that the number of actuators M and sensors S is identical, cf. Fig. 1.

The main task of the control is to set the desired

Machine speed (fixed setpoint control of the rigid body speed) and the suppression of local speed fluctuations (active

Torsional vibration damping). In this case, the position-synchronous movement of the individual print-image-carrying and sheet-transporting parts is of decisive importance for the print quality. The powertrain is an oscillatory system, which is essentially characterized by periodic, speed and

angle-dependent disturbances is excited to oscillations, whereby it comes to deviations in the position synchrony of individual parts, which are the

Can affect print quality negatively. In the design is on it Make sure that no tooth flank change occurs in the press bed, which would lead to significant position deviations and to an unusable print quality.

To model the essential torsional vibration behavior of the

Drivetrain in the lower frequency range up to 50 Hz are often used linear multi-body systems in which individual real mass inertias combined to inertial mass inertia m _k and a mechanical

Coordinate q _{k be} assigned. These inertias are by linear

Replacement rigidities k _k and replacement damping b _k connected to each other. Such a simplified, linear and passive multibody system with the

Force input vector f is determined by the equation of motion

1) with the mass matrix M, the damping matrix B and the stiffness matrix K. Often can be used as the basic form of the drive train

Torsionsschwinger with smooth shaft strand without branching and meshing accept. Fig. 2 shows the simplified further used here

Driveline model with N = 10 mechanical degrees of freedom and realistic parameters for a press module with ten modules.

The low and distributed via the drive train mechanical damping can be neglected for modal investigation. The damping matrix B from (equation 1) can be dispensed with, thus producing the undamped system M q + K q = f. Due to the symmetry of the mechanical system matrices M and K, a regular modal matrix Ψ ₀ ε R ^NxN can be found, with the aid of which the undamped system can be represented in the mod coordinate representation

Ι + Λ ₀ ρ = Ψ ₀ ^τ ί with q = 'lO p (equation 2) can be transformed with the diagonal matrix of the eigenvalues Λ ₀ ("mass normalization") The individual N differential equations of the second order in (Eq ) are decoupled and each describe an undamped PT2 oscillator, which is independent of the remaining PT2 oscillators, the N with respect to the mass and stiffness matrix orthogonal columns of the Modalmatrix Ψ ₀ are called eigenmodes ψ ₀ and describe, on the one hand, how the eigenmodes of the modal PT2 oscillators are put together and on the other hand they show the influence of the force vector f on the individual modal PT2 oscillators. It thus becomes clear that the k-th element of the i-th eigenform is a measure of the controllability and observability of the respective i-th unattenuated eigenvalue at the / c-th mechanical degree of freedom q _k . FIG. 3 shows the first three elastic eigenmodes of the exemplary system treated here without damping. Characteristic of the here examined unbesselte system are the

Antinodes, which - especially for low-frequency eigenvectors - each occur at the edge of the series oscillator (see Homogeneous

Torsionsschwingerkette).

The influence of a periodic disturbance excitation zj on a system output, such as the differential position d _{t of} two mechanical degrees of freedom, is determined in the frequency spectrum by the frequency response / i _{d £ <} _ _ZJ - (ω) with the short notation

{di <- Zj described. Different frequency responses can be in one

Frequency response matrix are summarized, such as the

Transmission paths from all disturbance inputs to all difference positions Η _ά ^ _ζ (ω) with the short notation (d <- z).

For the simple visual evaluation of many relevant frequency responses, the course of the Frobenius norm of the frequency response matrix over the frequency ω,

be used. This quadratic sum of all amplitude responses at the respective frequency location is shown in FIG. 4 for the example model with logarithmic amplitude used here (bold line). In addition, some amplitude responses from the frequency response matrix (thin lines) are shown for comparison. The scalar function of the Frobenius norm over the frequency leaves all

Recognize resonance peaks, which are in the individual Only partially show amplitude responses clearly, making the Frobenius standard well suited for the visual evaluation of a frequency response matrix. Characteristic for the force (difference) position amplitude response is the principal drop in the amplitude response to higher frequencies. The H2 controller standard will be used for later controller design

used in the vividly the function of the Frobenius standard on the

Frequency is integrated to a scalar size.

Control variables:

If several motors are used for the regulation, then the question arises, which local situation for the individual engines for the steaming of

Resonance peaks is favorable. From Fig. 3 it can be seen that in the mode shapes - especially for low-frequency resonances - always

Forming antinodes at the edges, resulting in a good

Obtain observability and controllability of the corresponding eigenvalue. The motivated positioning of two motors at the edges of the series oscillator is a disadvantage as opposed to a possible lifting of tooth flanks in the gear train. This may be due to disturbances that the

Temporarily changing the torque directions within the gear train - leading to the problems already mentioned. The approach followed in the following provides that the first drive is assigned to one end, preferably module one, and the remaining motors are distributed in such a way that the respective motor always drives beyond a possible following motor. Fig. 5 outlines the idea of an example in which the required drive torque is fed via two motors. To illustrate, the individual modules need an imaginary torque requirement of 1 each. The first motor in the first module (left) drives the following six modules beyond the feed point of the second motor in module five. Thus, the torque direction from left to right - depending on the torque distribution - insensitive Disturbers impulses. The positioning of a motor in the first module also provides the opportunity to compensate for an influential deterministic disturbance torque directly at the point of origin, which is provided by mechanisms for

Acceleration of the sheet on printing speed (machine speed) arises.

In turning machines exists, usually in the middle range of

Printing press, a turning unit, which also represents a comparatively large parasitic. An engine in the vicinity of the turning unit is also advantageous here to compensate for disturbances directly at the point of origin. In addition there is the requirement of a special Einrichtbetriebs, with which the

Printing press must be positioned very accurately at the location of the turning unit, for which an actuator in the immediate vicinity is also favorable. Control variables:

The rule goal is the most synchronous movement possible of all parts of the

Powertrain. In this case, a maximum of so many controlled variables can be selected as independent manipulated variables exist. For the control of the engine speed, the average speed variable v _ges can be used as a controlled variable, which from the pulse

is derived. As a result of this weighted averaging of the individual rotational speeds with the respective mass inertia, the controlled variable becomes largely independent of local deviations, in particular due to other proper motions.

As further control variables, the differences between the manipulated variables are proposed. With their help, the distribution of the drive torque can be influenced, responds to a change in the overall torque requirement of the printing press and thus a possible flank lift in the gear train be counteracted. In addition, the change of static difference angles between the individual modules can be counteracted (due to changed torque to be transmitted within the gear train) (circumferential register drift). Such a change may occur, for example, due to a changed engine speed or a changed heating state of the printing press. In practice, such a high difference between the setting torques is selected that occurring disturbances do not cause any significant changes in the torque flow and in no case a change of sign.

Measured variables:

For the selection of the measuring locations, the ends of the gear train are again favorable, analogous to the manipulated variables. In addition, for practical reasons, the setting locations themselves are advantageous. As feedback variables for the scheme will be

Speeds selected, since when using angular positions high

There are requirements for synchronous recording and angular resolution. In addition, a static speed feedback u = R q - with the controller matrix R and the resultant force vector f = u + z from manipulated variables u and interferences z - directly influences the damping matrix in the

Equation of motion (equation 1),

M q + (B - R) q + K q = z, (equation 6) and thus the system damping essential. A return of angular positions q would act directly on the stiffness matrix and mainly affect the frequencies ω _{ο ί} and the _eigenmodes ψ _{ο έ} . In practice, not all feedback variables used need to be measured, but missing variables can be generated with observer structures and thus sensor technology can be saved.

The state of the art is to use a single-loop PI controller structure for the speed control, which adjusts the setting torque via a

Main motor feeds in one place of the gear train. The I element in the controller ensures the stationary accuracy of the fixed setpoint control and the P element increases the dynamics of the control loop. If m control and n measurement outputs are available for the control, then a PI multivalue output feedback can be used, as shown in FIG. 6, in which a plurality of controlled variables y _R can be defined and predetermined via the command variables w. Overall, as many controlled variables can be defined as independent set inputs exist. Based on the

einschleifigen Pl controller affects here the output error e _y proportionally acting regulator matrix R ε m. ^mxn mainly due to the fast dynamics of the closed loop, whereas the controller matrix R, ε M ^mxm with the integrated control error e _R affects the slower control dynamics and analogous to the single-loop Pl controller for the stationary accuracy of the

Controlled variables ensures. The output error e _y describes the deviation between the (measured) system outputs y and the stationary end values of the design model, which with the precontrol matrix M _y ε m. ^nxm from the

Command value w are calculated. With the pre-control ^matrix M _u ε M ^mxm , the guiding ^behavior can be optimized, which ^subsequently - due to the comparatively slow setpoint change comparable to the oscillatable eigenvalues - is of little relevance and is not considered further (M _u = 0). In this structure, the machine _speed v _ges from (equation 5) and m-1 manipulated variable differences u _dk _ ₁ = u _dl -u _dk for k = 2... M are used as controlled variable. Thus, on the one hand, the stationary accuracy of the engine speed is ensured as well as the setting of the desired average torque differences allows, with a tooth flank lifting can be avoided.

As an alternative to the Pl multivoltage controller, the control structure in FIG. 7

proposed. Here, only the speed variable v _{ges is defined} as a controlled variable y _R , which is fed back to the system via an I-controller and an input coupling te E ^m . Using the vector t, a (fixed) percentage

Torque distribution can be achieved. In the stationary final state of the

Output error e _y (speed error, since y = q selected) to zero and the manipulated variable consists of only the I-part and the pilot control, and the percentage distribution to the control inputs is determined by t. The advantage here is a simpler structure, since only one I-element with a scalar parameter exists. If the required control loop dynamics for the engine speed significantly slower than the fast part of the closed loop, ie the damped natural vibrations is - which is true here - the machine speed control can be largely independent of R with the parameter r, set, which represents a further advantage.

Finally, the percentage distribution of the manipulated variable compared to the specification of differential moments may also be advantageous. A disadvantage is the fixed

Torque distribution is because formally the controller would have to be redesigned for each desired division. However, as will be discussed later, changing the layout without a new design is practically possible and, in practice, proves to be robustly stable.

The controller matrices of the presented structures are determined by minimizing a system standard. To describe this approach, initially only the static feedback via the matrix R is considered, ie the control target of the active vibration damping.

The aim is to minimize the positional deviations within the machine, which are caused by the disturbances on the individual masses described above. The effect of a periodic disturbance on a dynamic system is described by the frequency response of the system from the disturbance input to the considered output. It is therefore a reasonable control objective to minimize the frequency responses of all N noise inputs to all N- 1 position deviations in a suitable form. For this purpose, the H2 system standard is suitable, which corresponds to the sum of the quadratic integrated individual frequency responses as shown in (Eq. Static output feedback:

The relevant frequency response matrix (d <- z) from the disturbance inputs z to the positional deviations d of the closed loop depends on the Controller parameters R ab (Rj = 0). In order to make sense of the optimization problem for finding the control parameters, the input quantities u _R must also be taken into account. Classically, this can be done in the form of an additional quality term, so that m _R is solved in (|| d <- + y ² · | u _R <- z || l) (equation 7). By varying γ, the weighting of the manipulated variable can be determined. A given limit of the manipulated variables, for example l | u _R <- z || ₂ <10, can be maintained by a suitable variation of γ. Frequently, when optimizing system standards, a frequency-dependent weighting is also performed by comparing the considered frequency responses with a filter || W · (d <- z) ||| be multiplied. The reduction of the frequency responses then takes place increasingly in the frequency ranges in which the weighting filter W (eo) assumes high values. As a result, the amplitudes of the higher

Resonances here naturally fall sharply, but this is not absolutely necessary.

Be, unlike shown in Fig. 7, not only speeds but all

State variables are reduced, then the optimization problem (equation 7)

closed solvable. It can be shown in this case that the static

Return is optimal, that is also with the use of a dynamic regulator no better result would be achieved. In this case, the resulting controller corresponds to the Ricatti controller known in the relevant literature.

If, however, the regulator is limited to a static feedback of all or even only a part of the rotational speeds, in general no closed solution of (equation 7) is possible, but instead a numerical optimization must be carried out. This can be done with the function Systune of the Control System Toolbox of the program package Matlab. This also allows a direct consideration of inequality constraints, so that the optimization problem rnin || d <- z || ₂ ud B. || u _R <- z || ₂ <10 (equation 8) can be considered. Pl output feedback with input coupling:

. For the calculation of the control parameters R and r, the structure in Figure 7 has the optimization problem corresponding to min (|| d <- z || ₂ + y _v ^■ "- zlü ud B. || u _R <- z || ₂

^Ri (equation 9)

<10, || , <- z || ₂ < _{T be} extended. The manipulated variables of the two controller components are limited separately. The control variable necessary for the control of the rigid body speed results directly from the friction and the specified acceleration in the

Startup. The beyond existing torque reserve of the engine can (taking into account the tooth flank lift) on u _R to

Vibration damping can be used. Thus, the quantity a _T does not represent a true degree of freedom, and the separate restriction makes later the comparison of the various concepts easier. In the quality criterion, the integrated control deviation x, the rigid body speed instead of the

Control deviation e _R itself rated, as it the low frequencies are weighted more. It turns out that the required regulation of the rigid body speed is much slower compared to the remaining dynamics of the system and only

insignificantly depends on R. Conversely, the remaining dynamics are only negligibly influenced by r. Thus (equation 9) can be subdivided into two steps, in which R is first calculated via (equation 8). This makes it possible first to design R with regard to a good vibration damping, and then r, for example, to be easily adjusted based on the position of the dominant pole. The overall behavior of the control loop should then be critical be checked. If, in individual cases, a higher dynamic of the rigid body speed is necessary, the controller design must be carried out in a step according to (equation 9).

It should also be noted that the structure given by t

Torque distribution is a system parameter. If this is changed, the system changes and thus formally a new controller design is necessary to ensure stability and control quality. Alternatively, for a fixed controller, a test for stability with respect to all meaningful t (sum of the elements of t gives one and no negative elements occur in t) can be made by μ analysis. Provided the dynamics of the controlled rigid body speed

(Engine speed) is sufficiently slow compared to the oscillatory eigenvalues of the system, it can be expected that the robust stability can be detected, and beyond also shows only a small influence of t on the control behavior of the vibration damping. This can be the

Torque distribution without a new controller design to be adjusted during operation of the scheme.

PI multivariable output feedback:

For the calculation of the control parameters R and R _{ the structure in Fig. 6 becomes the optimization problem

ud B. || u _R <- z || ₂ <10, || u, ^ z || ₂ <α _λ solved. It is envisaged that the integrated deviation, _v

Rigid body speed and the integrated deviation of the manipulated variable difference ^X I, AU are weighted separately, for this purpose different dynamics

to be able to influence.

For the following comparison and discussion of different

Loop configurations becomes the uniform set limit of || u _R <- z || ₂ <10 used. This corresponds approximately to a same manipulated variable use. The I components in the controller are adjusted so that there is a slow dynamic the control of the rigid body speed or the torque differences and thus no significant effect on the vibration damping takes place. For this reason, it is sufficient to consider only the static feedback R in the following.

8 shows the course of the Frobenius norm (equation 3) from the interference inputs z to the angular differences d for the optical evaluation of the achievable ones

Vibration damping when using one, two or N = 10 motors. In the legend, the achieved H2 system standard is given for comparison. The unregulated system (dashed line, narrow) clearly shows the weakly damped resonances. The reference here is the fully-actuated system (solid line, narrow) with one motor per degree of freedom, which is expected to dampen the system as much as possible. Already with a motor at the position one (solid line, wide), a considerable damping of the first resonances at approx. 50 rad / s, 120 rad / s and 180 rad / s can be achieved. In contrast, the addition of a second motor at position five at the first and third resonance helps only insignificantly (dashed line, wide). One reason for this is the weak influence on the first and third eigenvalues at position five, which is shown in FIG. 3. The influence on the second eigenvalue, on the other hand, is good and explains the greater attenuation here. If the last position N was selected for the second motor

(Dot-dash line, wide), so the achievable attenuation - as the

To expect self-forms - to be stronger. In Fig. 8 below, the standard of the fault inputs to the manipulated variable u _{R is} shown. The essential control effort takes place in the relevant frequency band between about 25 rad / s to about 300 rad / s ("50 Hz).

Fig. 9 shows the influence of the recirculated variables in the solid

Engine positions one and five. The reference represents the state feedback (solid line, narrow). Are not all states but only all

Speeds fed back (dashed line, wide), the system can be virtually as well damped, what to expect. If only collocated speeds are returned, ie the speeds at the setting positions, then the achievable falls System damping already considerably lower (dashed line, wide). The feedback of another speed at the position N (solid line, wide) gives a noticeable increase (at the expense of the second resonance) for the damping of the first and third resonance.

Two controller structures for speed control of a printing press were presented and an optimization - based method for the design of the

Controller parameters specified. The latter allows a systematic and objective evaluation of various possible drive configurations. This was discussed as an example. The results coincide with the

qualitative statements from the consideration of eigenforms.

As a controlled variable for the engine speed, for example, a single speed of a mechanical coordinate or an average value of different speed variables can be used. As suggested, it may be advantageous to weight the individual rpm values with the respective mass inertia at this coordinate (derivation from the pulse rate). The controlled variable for the machine speed is thus more independent of local

Speed fluctuations, since this averaging, especially at the present here low mechanical damping, the Starkörperdrehzahl associated eigenform and thus is largely orthogonal with respect to the mass matrix to the remaining eigenmodes.

The use of the differences of the manipulated variables as further controlled variables or the adjustable division of the stationary total manipulated variable allows a further positive influence on the printing press. This controlled variable can influence the direction of the torque and avoid a change during operation, thus preventing possible tooth flank lifting. If the overall torque requirement of the press changes due to changed machine conditions such as engine speed, engine warm-up or pressure module conditions, etc., the torque flow within the gear train and a changed average displacement between the individual changes Modules adjusts. This printing technology disadvantageous effect can also be counteracted with the specification of differential torques.

Local angular or rotational speed deviations due to attacking periodic disturbances, for example due to periodic acceleration mechanisms or transient disturbances, for example due to a change in the operating state of the printing machine, for example the production speed (engine speed), are determined by means of the active torsional vibration damping by the static

Speed feedback reached.

Using modal analysis (eigenmodes) and more

pressure-machine-specific constraints were given advantageous positions for measuring and actuators and examined various possibilities by way of example.

Possibilities for visual evaluation and comparison of different controllers have been specified and used for control parameter optimization. In the case of low dynamic requirements of the

Guide behavior for the machine speed can be the design of the static speed feedback to influence the relevant (dynamic)

Eigenverhaltens (Störübertragungsverhalten) done separately. For both

Controller structures have been set up the optimization problems for a one-step or two-step approach.

Claims

claims

Control of printing machines, in particular sheet-fed printing machines, with one or more main drive motors (M), which have a

Powertrain are connected,

characterized in that

a central control unit (ZR) is used for the regulation of a machine speed and an active damping of vibrations,

wherein multiple speeds of different positions of the drive train are used for the control,

wherein the speeds are detected with sensors (S) and / or estimated with observers.

Control according to claim 1,

characterized in that

Each speed acts on all main drive motors or more

Ausgansfehler act on all main propulsion engines.

Control according to one or more of claims 1 to 2,

characterized in that

calculating a machine speed control variable from the weighted averaging of different current speeds,

the respective weights are determined by the mass inertia of the

Drivetrain result.

Control according to one or more of claims 1 to 3,

characterized in that

in addition to a machine speed control variable further control variables such as preferably adjusting torque differences can be selected.

Control according to one or more of claims 1 to 4,

characterized in that

as a controller structure, a PI multivoltage controller is used,

wherein the controller structure simplified with an input coupling such is realized that a fixed or variable depending on a state of the printing press torque distribution.

6. Control according to one or more of claims 1 to 5,

characterized in that

the differential torques are selected so that a tooth flank lift is prevented in the drive train.

7. Control according to one or more of claims 1 to 6,

characterized in that

the controller design is separated according to guiding behavior and interference behavior. 8. Control according to one or more of claims 1 to 7,

characterized in that

Controller parameters are obtained by optimizing a system standard, the system to be optimized has the positions of significant sources of interference as system inputs and the differential angle relevant to the print quality as system outputs.

9. Control according to one or more of claims 1 to 8,

characterized in that

be used for the positioning of the drive motors intrinsic forms of the drive train and positions of significant sources of interference,

wherein preferably the drive motors in the edge positions in the last and / or in the first module,

and / or in the area of a turning module.